Centrifugal compressor achieving high pressure ratio

ABSTRACT

Compressors achieve in a single stage a high-pressure ratio (r) of greater than or equal to 2.5:1 on a process fluid having a molecular weight of 12-20, such as natural gas. Two or more of the compressor stages are combined serially to increase overall pressure ratio. Each single-stage includes respective inlet and outlet passages and an unshrouded, centrifugal impeller that includes a plurality of impeller blades. Process fluid is discharged from trailing edges of the impeller blades at a rotational velocity greater than or equal to 1400 feet/second into a diffuser passage of the outlet. Dimension ranges of the annular diffuser passage, the centrifugal impeller, and the diffuser vanes vary as a function of pressure ratio (r) and/or the flow coefficient (φ) of the process fluid flowing between the inlet and the outlet.

PRIORITY CLAIM

This application is a continuation-in-part of U.S. utility patent application Ser. No. 14/272,667, filed May 8, 2014, and entitled “Supersonic Compressor”, which claims the benefit of U.S. Provisional Application No. 61/823,237, filed May 14, 2013, and entitled “Supersonic Compressor”, both of which are incorporated by reference herein. Priority under the parent applications is claimed in all jurisdictions where it is permissible to do so.

TECHNICAL FIELD

The invention relates to centrifugal compressors. More particularly, in some embodiments, the invention relates to centrifugal compressors and methods of their operation, which in a single compressor stage, is capable of achieving high-pressure ratios of greater than, or equal to 2.5:1 on process fluids having a molecular weight range of 12-20, such as natural gas. In other embodiments, the invention relates to two-stage centrifugal compressors and methods of their operation, which are capable of achieving throughput pressure ratios of greater than or equal to 5:1 on process fluids having a molecular weight of 12-20, such as natural gas.

BACKGROUND

Reliable and efficient compressors and systems including compressors have been developed and are utilized in a myriad of industrial processes (e.g., petroleum refineries, offshore oil production platforms, and subsea-process control systems). Generally, conventional compressors are utilized to compress gas or gas/liquid mixture process fluids, which are also referred to as “working fluids”. Typically, compression is achieved by applying mechanical energy to the process fluid gas in a low-pressure environment and transporting the gas to and compressing the gas within a high-pressure environment, such that the compressed gas may be utilized to perform work or for operation of one or more downstream process components.

As conventional compressors are increasingly used in offshore oil production facilities and other environments facing space constraints, there is an ever-increasing demand for smaller, lighter, and more compact compressors. In addition to the foregoing, it is desirable for commercial purposes that the compact compressors achieve higher overall throughput compression ratios (e.g., 10:1 or greater) while maintaining a compact arrangement.

In the past, higher compression ratios were achieved by increasing the number of compression stages within the compressor. Increasing the number of compression stages, however, increases the overall number of components (e.g., impellers and/or other intricate parts) required to achieve the desired compressor throughput (e.g., mass flow) and pressure rise to achieve the higher compression ratios. Increasing the number of components required in these compact compressors may often increase length requirements for the rotary shaft and/or increase distance requirements between rotary shaft bearings, leading to mechanical issues. The imposition of these requirements often results in larger, less compact compressors as compared to compact compressors utilizing fewer compression stages. Further, in many cases, increasing the number of compression stages in the compact compressors may still not provide the desired higher compression ratios or, if the desired compression ratios are achieved, the compact compressors may exhibit decreased efficiencies that make the compact compressors commercially undesirable. For example, compression ratio increase within a compressor stage of given inlet, impeller, diffuser passage, and outlet dimensions vary with molecular weight (mole weight) of the compressed process fluid. A compressor that is structurally configured to provide a pressure ratio of 11:1 for a 44 mole weight process fluid for a given impeller tip speed might only achieve a pressure ratio of 1.5:1 for a 10 mole weight process fluid. Generally, increasing impeller tip speed increases pressure ratio of the process fluid, up to its aero-thermodynamic and mechanical limits; i.e., Mach numbers for aero-thermodynamics and stress levels and material properties for mechanical.

At least one known proposed solution to the above-mentioned constraints of conventional compact compressors has been the utilization of supersonic compressors to achieve higher compression ratios while maintaining a compact structure. At least some of the known supersonic compressors utilize a compressor rotor that imparts supersonic velocity, greater than Mach 1, on the process fluid, to achieve greater single-stage pressure ratios than conventional compressors that impart velocities less than Mach 1.

SUMMARY OF INVENTION

Exemplary compressor embodiments described herein achieve high-pressure ratios of at least 2.5:1 in a single compressor stage, on process fluids having molecular weights of 12-20. Exemplary process or working fluids within the 12-20 molecular weight range include natural gas, comprising methane, and other hydrocarbons, with or without other non-hydrocarbon constituents. In other embodiments, the compressor is configured to impart a pressure ratio of at least 5:1 on the process fluid having a molecular weight of 24-27.99; or at least 4:1 on the process fluid having a molecular weight of 20-24; or at least 3:1 on the process fluid having a molecular weight of 16-20; or at least 2.5:1 on the process fluid having a molecular weight of 10-16; or at least 2:1 on the process fluid having a molecular weight less than 10. Other exemplary compressor embodiments described herein include first and second single-stage compressors serially in communication within a common housing structure or as separate housing structures, where both stages are commonly driven by a driver, such as an electric motor or a turbine engine. In such two-stage compressors, the outlet of the first-stage compressor, downstream of its diffuser passage, is in fluid communication with the inlet of the second-stage compressor. Thus, a multi-stage compressor of four stages, constructed and operated in accordance with embodiments described herein, is capable of compressing natural gas within a mole weight range of 12-20 to a commercially desirable pressure ratio of 39:1 or higher. Embodiments of the compressor stages described herein are of modular structure, which achieve maximum pressure ratios within a range of between 2.0:1-11:1 proportionally for 10-44 mole weight (MW) process fluids.

Each single-stage compressor, constructed in accordance with embodiments described herein, includes a housing having an inlet defining an inlet passage, and an outlet defining an annular diffuser passage. The process fluid in the inlet passage has an inlet pressure (P₁) and the process fluid discharged from the annular diffuser passage has a discharge pressure (P₂) greater than the inlet pressure, such that a pressure ratio (r) of the discharge pressure divided by the inlet pressure is greater than unity. A shaft-mounted centrifugal impeller is oriented between the inlet and the outlet. The impeller includes a plurality of three-dimensional impeller blades projecting outwardly from a hub. The hub has an axial length (A_(x)), and a hub outer diameter (D₂) extending radially at a radius (R₂) relative to the shaft axis. Each of the impeller blades has a leading edge facing the inlet passage at a blade sweep angle (θ), a trailing edge facing the annular diffuser passage at a back sweep angle (β) and having a tip width (b₂), and a blade tip having a radius of curvature (R_(C)), which defines an outer periphery of the centrifugal impeller. The impeller blades are configured to impart energy to the process fluid, upon rotation of the rotary shaft, and discharge the process fluid therefrom at a flow angle (α) into the annular diffuser passage. The annular diffuser passage has a leading or shroud wall and a trailing or hub wall, which define a diffuser passage height (b₃) there between. A plurality of diffuser vanes is oriented in the annular diffuser passage. The respective diffuser vanes extend axially from the shroud wall towards the hub wall of the diffuser passage. The diffuser vanes have a vane height (b_(3R)), and they are circumferentially disposed about the periphery of the centrifugal impeller. Each diffuser vane respectively defines curved, opposing vane pressure and vane suction sides, a vane leading edge proximate the periphery of the centrifugal impeller at a radial distance (R₃) relative to the shaft axis and conjoining the suction side, a vane trailing edge facing the outlet and conjoining the suction side, and a vane radial extent between the vane leading and trailing edges. The vane radial extent defines a length (RE). Dimension ranges of the annular diffuser passage, the centrifugal impeller, and the diffuser vanes vary as a function of pressure ratio (r). In various embodiments, when r varies between 2:1-10:1, A_(x)/D₂ varies between 0.1-0.4; 1/R_(C) varies between 0.5-0.15; 0; and 0 varies between 40-86. When r varies between 2:1-10:1, the angular bandwidth of minimum and maximum angles for the back sweep angle β vary between 20-60 degrees and 35-45 degrees, and the ratio of RE/R₂ varies from 0.1-0.5. Dimension ranges of the annular diffuser passage and the diffuser vanes vary as a function of flow coefficient (φ) of the process fluid flowing between the inlet and the outlet of the compressor. When φ is in the range of 0-0.030, the ratio of b_(3R)/b₃ is 1.0. When φ is in the range of 0.030-0.050, the ratio of b_(3R)/b₃ is 0.5-1.0. When φ is in the range of 0.050-0.110, the ratio of b_(3R)/b₃ is 0.3. Aforementioned dimensions of the modular construction housing, impeller, diffuser passage, diffuser vanes and outlet are selectively matched, within the aforementioned ranges, to achieve desired pressure ratios for given molecular weight process fluids. Achievable pressure ratios within any given stage are limited by ultimate mechanical stress limit of the impeller.

Other exemplary embodiments of the invention feature methods for compressing natural gas, having a molecular weight (MW) of 12-20. The methods are practiced by utilizing a single-stage, centrifugal compressor with a compressor casing having therein an inlet for receipt of a process fluid. The single-stage compressor includes a single, unshrouded, rotatable, centrifugal impeller defining a plurality of impeller blades for imparting kinetic energy into the process fluid. Each of the respective impeller blades has a leading edge for receiving process fluid from the inlet and a trailing edge for discharging process fluid. The single-stage compressor also has a diffuser, which defines an annular diffuser passage, for receiving the process fluid discharged from the respective trailing edges of the impeller blades in the annular diffuser passage and increasing static pressure of the process fluid therein. The compressor has an outlet for receiving process fluid discharged from the annular diffuser passage. When practicing the method, a first process fluid, comprising natural gas, having a molecular weight (MW) of 12-20, is introduced into the inlet passage of the compressor at an inlet pressure (P₁). The centrifugal impeller is driven at a rotational speed (N), so that the trailing edges of the respective impeller blades achieve a rotational velocity (U₂) of greater than or equal to 1400 feet/second: this imparts kinetic energy into the first process fluid. The diffuser passage receives the first process fluid discharged by the trailing edges of the centrifugal impeller blades, which converts the kinetic energy imparted in the first process fluid by centrifugal impeller into a pressure increase. The first process fluid is discharged from the annular diffuser passage at a discharge pressure (P₂) greater than the inlet pressure (P₁) thereof, such that a pressure ratio (r) of the discharge pressure divided by the inlet pressure is greater than or equal to 2.5:1.

Other exemplary embodiments of the invention feature methods for compressing process fluids. The methods are practiced by utilizing a single-stage, centrifugal compressor with a compressor casing having therein an inlet for receipt of a process fluid. The single-stage compressor includes a single, unshrouded, rotatable, centrifugal impeller defining a plurality of impeller blades for imparting kinetic energy into the process fluid. Each of the respective impeller blades has a leading edge for receiving process fluid from the inlet and a trailing edge for discharging process fluid. The single-stage compressor also has a diffuser, which defines an annular diffuser passage, for receiving the process fluid discharged from the respective trailing edges of the impeller blades in the annular diffuser passage and increasing static pressure of the process fluid therein. The compressor has an outlet for receiving process fluid discharged from the annular diffuser passage. When practicing the method, a first process fluid, having a molecular weight (MW), is introduced into the inlet passage of the compressor at an inlet pressure (P₁). The centrifugal impeller is driven at a rotational speed (N), so that the trailing edges of the respective impeller blades achieve a rotational velocity (U₂) of greater than or equal to 1400 feet/second: this imparts kinetic energy into the first process fluid. The diffuser passage receives the first process fluid discharged by the trailing edges of the centrifugal impeller blades, which converts the kinetic energy imparted in the first process fluid by centrifugal impeller into a pressure increase. The first process fluid is discharged from the annular diffuser passage at a discharge pressure (P₂) greater than the inlet pressure (P₁) thereof, such that a pressure ratio (r) of the discharge pressure divided by the inlet pressure is:

at least 5:1, where the first process fluid has a molecular weight of 24-27.99;

or at least 4:1 where the first process fluid has a molecular weight of 20-24;

or at least 3:1 where the first process fluid has a molecular weight of 16-20;

or at least 2.5:1 where the first process fluid has a molecular weight of 10-16;

or at least 2:1 where the first process fluid has a molecular weight less than 10.

The respective features of the exemplary embodiments of the invention that are described herein may be applied jointly or severally in any combination or sub-combination.

BRIEF DESCRIPTION OF DRAWINGS

The exemplary embodiments of the invention are further described in the following detailed description in conjunction with the accompanying drawings, in which:

FIG. 1 is a schematic view of an exemplary compressor, operatively coupled to a driver, in accordance with an embodiment as described herein;

FIG. 2 is a quartered cross-sectional, perspective view of an exemplary, single-stage compressor, in accordance with another embodiment as described herein;

FIG. 3 is a quartered, axial cross-sectional view of an exemplary, multi-stage, center-hung compressor; in accordance with another embodiment as described herein;

FIG. 4 is an axial cross-sectional view of another exemplary single-stage compressor in accordance with another embodiment as described herein;

FIG. 5 is a front elevational or radial, cross-sectional view of the compressor of FIG. 4, taken along 5-5 thereof, showing a portion of its centrifugal impeller and a modular, annular static diffuser within its annular diffuser passage;

FIG. 6 is an alternate embodiment of the impeller of FIGS. 4 and 5;

FIGS. 7A-C, respectively are axial cross-sectional alternative embodiments of annular diffuser passages and diffuser vanes, where dimensional ranges b_(3R)/b₃ vary as a function of flow coefficient (φ) of the process fluid flowing between the inlet and the outlet of the compressor, in accordance with embodiments of the invention;

FIG. 8 is an axial elevational view of an exemplary impeller, in accordance with embodiments of the invention;

FIG. 9 shows graphically the interrelationship of compressor impeller speed (in feet/second) and compressor pressure ratio (r) for process fluids having molecular weights (also referred to as mole weight or MW) between 10 and 44;

FIG. 10 shows graphically the interrelationship of the ratio of axial length (A_(x)) and corresponding outer diameter (D₂) and compressor pressure ratio (r), of exemplary centrifugal impellers, in accordance with embodiments of the invention;

FIG. 11 shows graphically the interrelationship of curvature of the blade tip of the exemplary three-dimensional blades of the centrifugal impeller of FIG. 10 (expressed as the inverse of the radius of curvature, 1/R_(C)) and compressor pressure ratio (r), in accordance with embodiments of the invention;

FIG. 12 shows graphically the interrelationship of blade sweep angle (θ) of the leading edge of exemplary three-dimensional blades of the centrifugal impeller of FIG. 10 and compressor pressure ratio (r), in accordance with embodiments of the invention;

FIG. 13 shows graphically the interrelationship of the back sweep angle (β) of the trailing edge of exemplary three-dimensional blades of the centrifugal impeller of FIG. 10 and compressor pressure ratio (r), in accordance with embodiments of the invention;

FIG. 14 shows graphically the interrelationship of the ratio of the radius of the diffuser vane leading edge to the outer hub radius of the impeller hub (expressed as R₃/R₂) and the exit flow angle (α) of process fluid off the impeller, of the exemplary impeller and diffuser vane of FIGS. 7A and 7B, in accordance with embodiments of the invention;

FIG. 15 shows graphically the interrelationship of the exit flow angle (α) of process fluid off the impeller, of the exemplary impeller and diffuser vane of FIGS. 7A and 7B and the corresponding ratio of the length of the vane radial extent to the radius of the impeller hub (expressed as RE/R₂), in accordance with embodiments of the invention; and

FIG. 16 shows graphically the interrelationship of the ratio of the diffuser vane height and the diffuser passage height (expressed as b_(3R)/b₃) and the flow coefficient (φ) of the process fluid flowing between the inlet and the outlet of the compressor, of the exemplary annular diffuser of FIGS. 7A and 7B.

To facilitate understanding, identical reference numerals have been used, where possible, to designate identical elements that are common to the figures. The figures are not drawn to scale.

DESCRIPTION OF EMBODIMENTS

Exemplary embodiments of the invention are utilized in compressors, and methods of their operation. These exemplary embodiments achieve, in a single compressor stage, between the inlet and outlet of the stage, a high pressure ratio (r) of at least 5:1 on the process fluid having a molecular weight of 24-27.99; or at least 4:1 on a process fluid having a molecular weight of 20-24; or at least 3:1 on a process fluid having a molecular weight of 16-20 or at least 2.5:1 on a process fluid having a molecular weight of 10-16 or at least 2:1 on a process fluid having a molecular weight less than 10. By way of example, an exemplary embodiment achieves a pressure ratio (r) of greater than or equal to 2.5:1 on a process fluid having a molecular weight of 12-20, such as natural gas. Natural gas in that molecular weight range typically comprises methane, other hydrocarbons, and non-hydrocarbon constituents, such as water, and carbon dioxide. In at least one embodiment, the process fluids pressurized, circulated, contained, or otherwise utilized in the supersonic compression system are a liquid phase, a gas phase, a supercritical state, a subcritical state, or any combination thereof.

Other exemplary compressor embodiments described herein include first and second single-stage compressors, serially in communication within a common housing structure or as separate housing structures, where both stages are commonly driven by a driver, such as an electric motor or a turbine engine. In such two-stage compressors, the outlet of the first-stage compressor, downstream of its diffuser passage, is in fluid communication with the inlet of the second-stage compressor. In compressor embodiments having more than two stages, the inlet of each successive stage is downstream of the outlet of the prior stage. In some embodiments, one or more compressor stages incorporate known intercooling structure, for extracting heat from the process fluid.

In some embodiments, the compressors are modular compressors, with a plurality of modular housings, respectively including annular diffuser passages and diffuser vanes; and a plurality of modular impellers. Internal dimensions of the respective modular housings and impellers are matched for imparting desired pressure ratios in selected process fluids respectively having varying molecular weight range properties during compressor design. In some embodiments, the diffuser vanes are also modular, annular diffuser vanes having varying vane height. Exemplary internal dimensions of the modular housings, modular diffuser vanes, and the modular centrifugal impellers are described in detail herein.

Compressor Structure

FIGS. 1 and 2 illustrate, a schematic view of an exemplary compression system 100 including a compressor 102, according to one or more embodiments. During compression, the compressor 102 imparts a supersonic velocity (i.e., greater than or equal to Mach 1) on a process or working fluid to increase the compressor's pressure ratio. Hence, the compressor 102 is alternatively referred to herein as a supersonic compressor. A driver 104, such as any one or more of an electric motor, hydraulic motor, internal combustion engine or a turbine engine of known construction, is operatively coupled to the supersonic compressor 102 via a drive shaft 106. In exemplary alternative embodiments, the drive shaft is integral with or coupled with a rotary shaft 108 of the supersonic compressor 102. Drive shaft 106 is coupled with the rotary shaft 108 via a gearbox 109 of known construction, including a plurality of gears configured to transmit the rotational energy of the drive shaft 106 to the rotary shaft 108 of the supersonic compressor 102, such that the drive shaft 106 and the rotary shaft 108 alternatively spin at the same speed, substantially similar speeds, or disparate speeds. The rotary shaft 108 spins about its central rotational axis 108A.

In the exemplary embodiments of FIGS. 2 and 4-6, the compressor 102 is a direct-inlet, or axial-inlet, single-stage, centrifugal compressor, with an overhung rotor configuration. Alternatively, as shown in FIG. 3, the compressor 102A is a multi-stage, center-hung compressor, having a radial inlet. Three stages, referenced as SI, SII, and SIII, are substantially similar in construction to the single stage of the compressor 102. For practical purposes in understanding, structure and operation of the inventions described herein, each of the respective stages SI, SII, and SIII is constructed and functions in a substantially similar manner as the single stage of the compressor 102. In multi-stage embodiments, whether two-stage or greater than two-stage, the inlet of each successive stage is in communication with the outlet of the prior stage. Thus, the outlet of SI is upstream of and is in direct fluid communication with the inlet of SII, and the outlet of SII is upstream of and is in direct fluid communication with the inlet of SIII. Pressurized throughput of process fluid in the compressor 102A is from the inlet 102B of SI to the outlet 102C of SIII. In some embodiments, one or more stages SI-S_(N), where N equals the total number of stages in the compressor, are cooled by a known intercooler(s) (not shown). In view of the structural and functional similarities of each single stage SI-SIII of the motor 102A, further description of an exemplary compressor stage is focused on the single-stage compressor 102.

Referring to FIGS. 2 and 4-7A, B, C, the single-stage compressor 102 embodiments respectively include a housing 120 having an inlet 122 defining an inlet passage 124, and an outlet 130 defining an annular diffuser passage 140. The process fluid in the inlet passage 124 has an inlet pressure P₁ and the process fluid discharged from the diffuser passage 140 into the outlet 130 has a discharge pressure P₂ greater than the inlet pressure, such that a pressure ratio r of the discharge pressure divided by the inlet pressure is greater than unity. A rotary shaft 108, defining a shaft axis 108A, is oriented in the housing 120 between the inlet and the outlet thereof, driven by the driver 104. In some applications, one or more optional inlet guide vanes 126 guide and direct incoming process fluid flow a centrifugal impeller 160. When utilized, the inlet guide vanes 126 are configured to condition the process fluid flowing therethrough to include one or more predetermined parameters, such as a circumferential swirl, a velocity, a mass flow rate, a pressure, a temperature, and/or any suitable flow parameter to enable the supersonic compressor 102 to function as described herein. In some embodiments, the inlet guide vanes 126 are static or they are adjustable. In an exemplary embodiment, a plurality of inlet guide vanes 126 are arranged about a circumferential inner surface of the inlet 122 in a spaced apart orientation. The spacing of the inlet guide vanes 126 may be equidistant or may vary depending on the predetermined, desired flow conditioning parameter for the process fluid.

As shown in FIG. 2, the shaft-mounted centrifugal impeller 160 is oriented between the inlet 122 and the outlet 130. Referring to FIGS. 7-9, the impeller 160 includes a plurality of three-dimensional impeller blades 162 projecting outwardly from a hub 164. The hub 164 is concentric with the rotational axis 108A of the shaft 108. The hub 164 has an axial length A_(x), and a hub outer diameter D₂ extending radially at a radius R₂ relative to the rotational axis 108A. The each of the impeller blades 162 has: a leading edge 166 facing the inlet passage 124 at a blade sweep angle θ; a trailing edge 168 facing the diffuser passage 140 at a back sweep angle β, and having a tip width b₂; and a blade tip 170 having a radius of curvature R_(C), which defines an outer periphery of the centrifugal impeller. The radius of curvature R_(C) is measured along a circle 171 that is coplanar with the leading edges 166 of the blades 162; the circle is concentric with the rotational axis 108A, and has a diameter that is equal to the diameter D₂ of hub 164. The hub diameter D₂ is equal to twice the radius of the hub R₂.

In an exemplary embodiment of FIG. 6, the plurality of blades 262 include one or more splitter blades 263, configured to reduce choking conditions that may occur in the supersonic compressor 102 depending on the number of blades 262 employed with respect to the centrifugal impeller 260. A splitter blade 263 includes a leading edge 263A that is not coplanar with at least one other leading edge 262A of the centrifugal impeller 260.

Referring to FIGS. 5 and 8, the impeller blades 162 are curved in three dimensions, such that upon rotation of the shaft 108, the process fluid flowing into the inlet 122 is drawn into the centrifugal impeller 160 and accelerated in a tangential and radial direction by the centrifugal force imparted by the impeller 160. The accelerated process fluid is discharged from the blade trailing edges 168 in radial directions that extend 360 degrees around the periphery of the centrifugal impeller, thereby increasing the velocity and static pressure of the process fluid. In an embodiment, the velocity of the process fluid at the trailing edges 168 of the blades 162 is about Mach 1 or greater. In other embodiments, the velocity of the process fluid at the same trailing edges 168 position is a between about Mach 1.5 and about Mach 3.5, although wider ranges are certainly possible within the teachings hereof, within strength limits of material forming the impeller 160 and geometric constraints of the impeller. The process fluid is discharged from the trailing edges 168 at a flow angle α into the diffuser passage 140, as indicated by the arrow F.

The centrifugal impeller 160 embodiments shown in all of the figures are open or “unshrouded”, because they do not incorporate a rotating shroud that defines a boundary of the process fluid path between the inlet 122 and the outlet 130 upstream of the centrifugal impeller. Unshrouded impellers are capable of achieving higher rotational speeds of the blade tips than shrouded designs, enabling higher compression ratios for any given process fluid molecular weight range. In multi-stage compressor embodiments, such as the compressor 102A of FIG. 3, each of the impellers is unshrouded. In the embodiment of FIG. 4, the boundary of the process fluid path between the inlet 122 and the outlet 130 upstream of the centrifugal impeller (i.e. on the left side of the figure) is defined by the wall of the housing 120, and is referred to as the “shroud side” of the housing. The boundary of the process fluid path between the inlet 122 and the outlet 130 downstream of the centrifugal impeller (i.e. on the right side of the figure) is defined by the wall of the housing 120, and is referred to as the “hub side” of the housing. In other embodiments, the centrifugal impeller is semi-open or shrouded.

As shown in FIG. 5, the process fluid is discharged from the trailing edges 168 of respective blades 162 of the impeller 160 at the flow angle α into the annular diffuser passage 140. Referring to FIGS. 5 and 7A-7C, the annular diffuser passage 140 has a shroud wall 142 and a hub wall 144, which define the diffuser-passage height b₃ there between. Though not drawn to scale in FIGS. 7A-C, there is a diffuser pinch in the diffuser passage 140, so that the diffuser-passage height b₃ is less than the tip width b₂ of the impeller trailing edge 168. In the embodiments of FIGS. 7A and 7B, a plurality of diffuser vanes 146 is oriented in the annular diffuser passage 140. The respective diffuser vanes 146 convert kinetic energy imparted in the process fluid by the impeller 160 into static pressure increase, which increases the process fluid's potential energy. The diffuser vanes 146 extend axially from the shroud wall 142 towards the hub wall 144 of the diffuser passage 140. The diffuser vanes 146 have a vane height b_(3R), and are circumferentially disposed about the periphery of the centrifugal impeller 160. In FIGS. 4, 5 and 7B, the vane height b_(3R) of the vanes 146 spans the entire passage height b₃ of the diffuser, such that the ratio b_(3R)/b₃ equals unity.

In some embodiments, each diffuser vane 146, regardless of its vane height b_(3R), respectively defines curved, opposing vane pressure 148 and vane suction sides 150, a vane leading edge 152 proximate the periphery of the centrifugal impeller at a radial distance R₃ relative to the shaft axis 108A and conjoining the suction side 150, a vane trailing edge 154 facing the outlet 130 and conjoining the suction side 150. A vane radial extent, between the vane leading 152 and trailing 154 edges, has a radial length RE measured relative to the shaft rotational axis 108A. In the vane 146 embodiment of FIG. 7A, the vane height b_(3R) does not extend all the way across the diffuser passage height b₃, such that the ratio b_(3R)/b₃ is less than unity. As previously mentioned, in the vane embodiment of FIG. 7B, the vane 146 spans the entire diffuser passage height. The embodiment of FIG. 7A incorporates a second vane 146A downstream of the vane 146. The second vane 146A has the same construction features as the vane 146 in FIGS. 7A and 7B. In FIG. 7B, when the vane 146 spans the entire diffuser passage height b₃, the second vane 146A is optional, as shown in phantom lines. In the embodiment of FIG. 7C, there are no vanes in the diffuser passage 140.

FIG. 5 illustrates a front, radial view taken along line 5-5 of a portion of the centrifugal impeller 160 and the annular diffuser passage 140 of FIG. 4. In the embodiment of FIGS. 4 and 5, the annular diffuser passage 140 incorporates a static diffuser, constructed as a modular, annular vane ring 155, which is inserted or into or interposed within the housing 120, between the impeller 160 and the outlet 130. Thus, vane 146 profiles within the annular diffuser passage 140 are easily varied by substitution of different annular vane rings 155 for different compressor applications.

As noted, the annular diffuser 140 is configured to convert kinetic energy of the process fluid from the centrifugal impeller 160 into increased static pressure. An annular-vane, static diffuser 140 is shown in the exemplary embodiments of FIGS. 4, 5, 7A, and 7B. In these exemplary embodiments, the plurality of diffuser vanes 146 controls the rate of change of, and thus the diffusion or area increase there between. Each of the diffuser passageways 156 is bounded radially by the respective pressure 148 and suction sides 150 of opposing diffuser vanes 146; and bounded axially by the respective shroud 142 and hub 144 walls of the annular diffuser passage 140.

In some embodiments, such as in FIG. 5, the diffuser passageway 156 defines a subsonic diffusion zone 157. In some embodiments, one or more of the plurality of diffuser vanes 146 includes supersonic compression-inducing surface at its leading edge 152, between its pressure side 148 and its suction side 150, such as the ramp 158 In an exemplary embodiment, each of the diffuser vanes 146 includes a supersonic, compression-inducing ramp 158 disposed at its leading edge 152. In some embodiments, the supersonic, compression-inducing ramp 158 is integral with the respective diffuser vane 146; however, in some embodiments, one or more of the supersonic compression-inducing ramps 158 are formed as a separate component. As the centrifugal impeller 160 is rotated, the radial process fluid flow F exiting the blades 162 at the trailing edges 168 enters each of the diffuser passageways 156 and exits the diffuser passageway via the respective diffuser passageway outlet 140.

Accordingly, the supersonic compressor 102 provided herein is said to be “supersonic” because the centrifugal impeller 160 designed to rotate about the shaft axis of rotation 108A at high speeds such that a moving process fluid encountering the supersonic compression-inducing surface 158 disposed within the diffuser passageway 156 is said to have a fluid velocity which is supersonic. Thus, in an exemplary embodiment, the moving process fluid encountering the supersonic compression-inducing surface 158 may have a velocity in excess of Mach 1. However, to increase total energy of the fluid system, the moving process fluid encountering the supersonic compression-inducing surface 158 may have a velocity in excess of Mach 1.2. The exemplary compression inducing surface 158 shown in FIG. 5 has a ramp shape.

Compressor Operation

FIG. 9 shows the physical challenge facing compressor designers who want to maximize single-stage pressure ratio r for different mole weight (MW) process fluids. Impeller tip speed needed to add energy to any given process fluid to a desired pressure ratio is directly proportional to mole weight. Achievement of supersonic velocities in the working fluid by impeller and/or annular vane structure enhances compression capabilities, as is done in embodiments of the invention disclosed herein. Despite achievement of greater than Mach 1, velocities in working fluid by embodiments of compressors disclosed herein, impeller tip speed is limited by mechanical stress limits of the impeller geometry and its material. Exemplary centrifugal impellers described herein are capable of achieving impeller tip speed rotational velocities at their blade trailing edges greater than or equal to 1400 feet/second. The horizontal stress limit line shown in FIG. 9 is in a typical range for titanium alloys. Lower mole weight fluids, e.g., in the 10 MW range, are not compressible in a single stage greater than about 2:1 with current impeller stress limits. Conversely, higher mole-weight process fluids are compressible to 9:1 or greater pressure ratios. As is shown in FIG. 9, conventional compressors that limit impeller tip speed and pressure ratio r to the lower left quadrant can achieve pressure ratios of approximately 1.05:1 to about 5:1 for a broad variety of process fluids, so long as the compressor user is willing to add more stages in the compressor to achieve desired pressure ratio throughput.

Embodiments of the present invention achieve higher-pressure ratios r for a broad range of disparate process fluids having widely varying mole weight ranges (MW). By utilizing combinations of compressor housings 120, centrifugal impellers 160 driven at blade tip speeds of greater than or equal to 1400 feet/second, annular diffusers 140, and annular vane rings 155, having specified differing dimensional ranges in accordance with embodiments of the invention, higher pressure ratios r and stage head are achieved by tailoring the compressor 102 structure to the MW properties of specific process fluids. Construction of the invention's housings, impellers, and discharge vane embodiments allows precise configuration and assembly of a specific compressor to meet needs of a narrow bandwidth of MW.

In embodiments disclosed herein, dimension ranges of the annular diffuser passage 140, the centrifugal impeller 160, and the diffuser vanes 146, 146A, and housings 120 needed to support those components vary as a function of pressure ratio r, regardless of the process fluid MW. In various embodiments, when r varies between 2:1-10:1, the impeller 160 dimensional ratio of hub length 164 to diameter A_(x)/D₂ varies between 0.1-0.4, as shown in FIG. 10. Higher-pressure ratios require axially longer impellers for a given impeller diameter. One way to reduce the impeller inventory is to standardize hub diameter D₂ and select a limited number of axial lengths A_(x). Concomitantly, the number of types of housing 120 modules only needs to match the limited number of impeller lengths.

As shown in FIG. 11, Curvature 1/R_(C) of the blade tip 170 of the impeller blades 162, varies between 0.5-0.15 for pressure ratios r between 2:1 and 10:1, regardless of the process fluid MW. As the process-fluid is accelerated between the leading edge 166 to the trailing edge 168 of the impeller blades 162 pressure imparted on the fluid increases, so blade curvature decreases. In other words, the radius R_(C) becomes steeper closer to the trailing edge 168 of the impeller blades 162. For modular impeller 160 construction simplicity, only one curvature profile is needed for each of the chosen modular impeller lengths A_(x). By limiting each impeller module 160 to a single blade curvature profile, it follows that only one modular housing 120 configuration is needed for each of the impeller modules.

In FIG. 12, the relationship of the sweep angle θ of the leading edge 166 of the impeller blades 162 varies between 40-86 or pressure ratios r between 2:1 and 10:1, regardless of the process fluid MW. Thus, in some embodiments, the sweep angle θ can be selectively matched to a given axial length A_(x) impeller module. Referring to FIG. 13, when r varies between 2:1-10:1, the angular bandwidth of minimum and maximum angles for the back sweep angle β of the trailing edge 168 of the impeller blades 162 vary between 20-60 degrees (at r=2:1) and 35-45 degrees (at r=10:1), regardless of the process fluid MW.

Impeller exit flow angle α is influenced by the impeller back sweep angle β, the impeller 160 tip speed at its trailing edge 168 and the mole weight of the process fluid. Impeller tip speed and the back sweep angle are selected to achieve a desired pressure ratio. Referring to FIG. 14, the ratio of the radius R₃ of the vane leading edge 152 to the outer hub radius R₂ (R₃/R₂) is 1.07-1.15, between a range of impeller exit flow angle α, of 68 to approximately 80 degrees. As shown in FIG. 15, the ratio of the diffuser vane 146 radial extent RE to the impeller hub radius R₂ (RE/R₂) varies approximately linearly from 0.1-0.5 between same range of impeller exit flow angles α, between 70 degrees and the maximum critical angle of 80 degrees.

Dimension ranges of the respective heights b_(3R) of the diffuser vanes 146 and b₃ of the annular diffuser passage 140 vary as a function of flow coefficient (φ) of the process fluid flowing between the inlet 122 and the outlet 130 of the compressor 102. Referring to FIG. 16, when φ is in the range of 0-0.030, the ratio of b_(3R)/b₃ is 1.0; the vane 146 spans the entire height of the annular diffuser passage 140. When φ is in the range of 0.030-0.050, the ratio of b_(3R)/b₃ is 0.5-1.0, and the vane 146 spans between half and the full height of the annular diffuser passage. Use of a second vane 146A is optional when φ is in the range of 0.030-0.050. When φ is in the range of 0.050-0.110, the ratio of b_(3R)/b₃ is 0.3.

Single-stage, centrifugal compressors constructed in accordance with embodiments described herein are capable of driving their respective, single, unshrouded impellers at rotational velocities of greater than or equal to 1400 feet/second. The high kinetic energy imparted by these impellers on the process fluid achieves the following pressure ratios (r):

at least 5:1, where the first process fluid has a molecular weight of 24-27.99;

or at least 4:1 where the first process fluid has a molecular weight of 20-24;

or at least 3:1 where the first process fluid has a molecular weight of 16-20;

or at least 2.5:1 where the first process fluid has a molecular weight of 10-16;

or at least 2:1 where the first process fluid has a molecular weight less than 10.

More particularly, commonly available natural gas compositions have molecular weight ranges of 12-20. Single-stage compressor embodiments disclosed herein, with tip speed velocities of greater than or equal to 1400 feet/second, are capable of achieving pressure ratios (r) greater than or equal to 2.5: 1 on such commonly available natural gas compositions in the 12-20 MW range.

In some embodiments, two of the aforementioned single-stage, centrifugal compressors are combined in series, with the outlet of the first stage coupled to the inlet of the second stage. The two stages in combination receive the process fluid in the inlet of the first stage, and discharge the process fluid out of the outlet of the second stage at a throughput pressure ratio (r) of:

at least 10:1, where the first process fluid has a molecular weight of 24-27.99;

or at least 8:1 where the first process fluid has a molecular weight of 20-24;

or at least 6:1 where the first process fluid has a molecular weight of 16-20;

or at least 5:1 where the first process fluid has a molecular weight of 10-16;

or at least 4:1 where the first process fluid has a molecular weight less than 10.

Two-stage, centrifugal compressor embodiments disclosed herein, with unshrouded impeller, tip speed velocities of greater than or equal to 1400 feet/second, are capable of achieving pressure ratios (r) greater than or equal to 5:1 on commonly available natural gas compositions in the 12-20 MW range.

In other embodiments, multi-stage compressors, having 3 or more of the aforementioned single-stage, centrifugal compressors in series, are configured to impart sequentially with the assembled multi-stage compressor a throughput pressure ratio (r) of greater than or equal to 10:1 on process fluids having a molecular weight of 2.0-27.99. Process fluids in the aforementioned molecular weight range include compositions of natural gas. Additionally, multi-stage compressors, having 3 or more of the disclosed, single-stage, centrifugal compressors in series, are configured to impart sequentially compressor a throughput pressure ratio (r) of greater than or equal to 10:1 on commonly available natural gas compositions in the 12-20 MW range.

Definitions

Terms used herein are defined as follows.

“Actual cubic feet per minute” (ACFM) is the volume of process gas flowing at the inlet to the compressor independent of its density. ACFM is related to the mass flow of the process fluid as follows:

${ACFM}{= \frac{WRT_{s}z_{s}}{144P_{s}M\; W}}$

Where R=the universal gas constant (1545.35 lb-ft/° F.-lbmmol)

-   -   MW=molecular weight of the process fluid     -   T_(s)=suction temperature at the compressor inlet, in ° R     -   z_(s)=compressibility of the working fluid at the compressor         inlet     -   P_(s)=absolute suction pressure at the compressor inlet         (referred to as P₁ in the drawings), in absolute pounds per         square inch (PSIA)     -   W=mass flow of the process fluid, in absolute pounds per minute         (Lb/Min)

-   “Flow coefficient” (φ) relates to an impeller's volumetric flow     capacity Q, in actual cubic feet per minute (ACFM), impeller     operating rotational speed N in feet/second, and the impeller hub     exit diameter D₂, in inches.

$\phi = \frac{Q}{N \times D_{2}^{3}}$

-   “Molecular weight” or “mole weight” (MW) is the sum of the atomic     weights of each constituent element multiplied by the number of     atoms of that element in its molecular formula. -   “Pressure ratio” (r) is the ratio of discharge pressure (P₂) of     process fluid discharged from the compressor outlet (downstream of     the compressor's annular diffuser passage) to inlet pressure (P₁). -   “Stage head” (H) is the measure of the amount of kinetic energy     required to elevate (in feet) a fixed amount of process fluid in a     given compressor stage by the pressure ratio r, from its inlet     pressure (P₁ or P_(s)) to its discharge pressure (P₂). The required     kinetic energy is related to the impeller tip speed of the impeller     blade trailing edges in accordance with the following equations.

$H = {\frac{n}{n - 1}{\frac{zRT_{s}}{g_{c}}\left\lbrack {\left( \frac{P_{2}}{P_{1}} \right)^{\frac{n - 1}{n}} - 1} \right\rbrack}}$

Where: n=the polytropic exponent

$\left( {\frac{n - 1}{n} = {\frac{1}{\eta}\frac{k - 1}{k}}} \right)$

-   -   k=the pseudo isentropic exponent     -   η=compressor stage polytropic efficiency     -   g_(c)=the gravitational Constant (32.174 ft/sec²)     -   R=the universal gas constant (1545.35 lb-ft/° F.-lbmmol)     -   T_(s)=suction temperature at the compressor inlet, in ° Rankine     -   z=compressibility of the process fluid         Stage head is related to impeller tip speed (U₂, in ft/second)         which is the rotational speed of the impeller blade trailing         edges as follows:

$H = \frac{\mu \; U_{2}^{2}}{g_{c}}$

Where: μ=the stage head coefficient (relationship of head increase, impeller rotational speed and impeller hub diameter)

-   -   g_(c)=the gravitational Constant (32.174 ft/sec²)

$U_{2} = \frac{\prod{D_{2}N}}{720}$

-   -   D₂=hub outer diameter of the impeller, in inches     -   N=impeller rotational speed, revolutions per minute         Alternatively, the stage head coefficient, μ, is expressed by         the following equation:

$\mu = \frac{\left\lbrack {\frac{n}{n - 1}zR{T_{S}\left\lbrack {\left( \frac{P_{2}}{P_{1}} \right)^{\frac{n - 1}{n}} - 1} \right\rbrack}} \right\rbrack g_{c}}{U_{2}^{2}}$

Although various embodiments that incorporate the invention have been shown and described in detail herein, others can readily devise many other varied embodiments that still incorporate the claimed invention. The invention is not limited in its application to the exemplary embodiment details of construction and the arrangement of components set forth in the description or illustrated in the drawings. The invention is capable of other embodiments and of being practiced or of being carried out in various ways. In addition, it is to be understood that the phraseology and terminology used herein is for the purpose of description and should not be regarded as limiting. The use of “including,” “comprising,” or “having” and variations thereof herein is meant to encompass the items listed thereafter and equivalents thereof as well as additional items. Unless specified or limited otherwise, the terms “mounted”, “connected”, “supported”, and “coupled” and variations thereof are used broadly and encompass direct and indirect mountings, connections, supports, and couplings. Further, “connected” and “coupled” are not restricted to physical, mechanical, or electrical connections or couplings. 

What is claimed is:
 1. A compressor comprising: a housing having: an inlet defining an inlet passage; and an outlet defining an annular diffuser passage, respectively configured to receive and flow a process fluid there between and therethrough, the process fluid in the inlet passage having an inlet pressure (P₁) and the process fluid discharged from the annular diffuser passage having a discharge pressure (P₂) greater than the inlet pressure, such that a pressure ratio (r) of the discharge pressure divided by the inlet pressure is greater than unity; a rotary shaft, defining a shaft axis, in the housing between the inlet and the outlet thereof; an unshrouded centrifugal impeller mounted about the rotary shaft between the inlet and the outlet of the housing, and in fluid communication with process fluid flowing there between, the centrifugal impeller having: a hub with an axial length (A_(x)) extending axially along the shaft axis of the rotary shaft, and a hub outer diameter (D₂) extending radially at a radius (R₂) relative to the shaft axis; a plurality of impeller blades projecting outwardly from the hub, each of the impeller blades having: a leading edge facing the inlet passage of the housing at a blade sweep angle (θ), a trailing edge facing the annular diffuser passage at a back sweep angle (β) and having a tip width (b₂), and a blade tip having a radius of curvature (R_(C)), which defines an outer periphery of the centrifugal impeller; the impeller blades configured to impart energy to the process fluid, upon rotation of the rotary shaft and, to discharge the process fluid therefrom at a flow angle (α) into the annular diffuser passage; the annular diffuser passage having: a shroud wall and a hub wall which define a diffuser passage height (b₃) there between; a plurality of diffuser vanes in the annular diffuser passage to receive process fluid discharged by the centrifugal impeller and convert the energy imparted therein by the centrifugal impeller, in order to raise pressure thereof by the pressure ratio and cause actual volumetric flow therethrough at a flow coefficient (φ), the respective diffuser vanes extending axially from the shroud wall towards the hub wall thereof and having a vane height (b_(3R)), and circumferentially disposed about the periphery of the centrifugal impeller, each diffuser vane respectively defining: curved, opposing vane pressure and vane suction sides, a vane leading edge proximate the periphery of the centrifugal impeller at a radial distance (R₃) relative to the shaft axis and conjoining the suction side, a vane trailing edge facing the outlet and conjoining the suction side, and a vane radial extent between the vane leading and trailing edges, the vane radial extent defining a length (RE); and wherein: the compressor is configured to impart a pressure ratio (r) of at least 5:1 on the process fluid having a molecular weight of 24-27.99; or at least 4:1 on the process fluid having a molecular weight of 20-24; or at least 3:1 on the process fluid having a molecular weight of 16-20; or at least 2.5:1 on the process fluid having a molecular weight of 10-16; or at least 2:1 on the process fluid having a molecular weight less than 10; the flow angle (α) is 69-80 degrees; ratio of the radius of the vane leading edge to the outer hub radius (R₃/R₂) is 1.07-1.15; and dimension ranges of the annular diffuser passage, the centrifugal impeller, and the diffuser vanes vary as a function of pressure ratio (r); and dimension ranges of the annular diffuser passage and the diffuser vanes vary as a function of flow coefficient (φ) of the process fluid flowing between the inlet and the outlet of the compressor as follows: r A_(x)/D₂ 1/R_(C) θ β (angular bandwidth) RE/R₂ 2:1-10:1 0.1-0.4 0.5-0.15 40-85 20-60 to 35-45 0.1-0.5

φ b_(3R)/b₃    0-0.030 1.0 0.030-0.050 0.5-1.0 0.050-0.110 0.3


2. The compressor of claim 1, comprising a single-stage compressor.
 3. A two-stage compressor, comprising two of the single-stage compressors of claim 2 in series, the stages in combination configured to impart a pressure ratio (r) of at least 10:1 on the process fluid having a molecular weight of 24-27.99; or at least 8:1 on the process fluid having a molecular weight of 20-24; or at least 6:1 on the process fluid having a molecular weight of 16-20; or at least 5:1 on the process fluid having a molecular weight of 10-16; or at least 4:1 on the process fluid having a molecular weight less than
 10. 4. The two-stage compressor of claim 3, configured to provide a pressure ratio (r) of greater than or equal to 5:1 on the process fluid having a molecular weight of 12-20, the process fluid comprising natural gas.
 5. The two-stage compressor of claim 3, further comprising the two single-stage compressors in series fluid communication, in a common in-line housing,
 6. The two-stage compressor of claim 5, configured to provide a pressure ratio (r) of greater than or equal to 5:1 on the process fluid having a molecular weight of 12-20, the process fluid comprising natural gas.
 7. The compressor of claim 1, the centrifugal impeller having at least one leading edge of an impeller blade that is not coplanar with the leading edge of at least one other impeller blade.
 8. The compressor of claim 1, further comprising a first row of a plurality of diffuser vanes; and a second row of a plurality of diffuser vanes whose respective second-row vane leading edges face the vane trailing edges of the diffuser vanes of the first row of said vanes, and whose respective second-row vane trailing edges face the outlet, and terminate at a radial distance (R₅) relative to the shaft axis.
 9. The compressor of claim 1, the diffuser vanes further comprising a compression-inducing surface formed along the vane leading edge, and configured to generate a shock wave from within the process fluid.
 10. A method for compressing natural gas, having a molecular weight (MW) of 12-20, comprising: providing a single-stage, centrifugal compressor with a compressor casing having therein: an inlet for receipt of a process fluid; a single, unshrouded, rotatable, centrifugal impeller defining a plurality of impeller blades for imparting kinetic energy into the process fluid, each of the respective impeller blades having a leading edge for receiving process fluid from the inlet and a trailing edge for discharging process fluid therefrom; a diffuser, defining an annular diffuser passage, for receiving the process fluid discharged from the respective trailing edges of the impeller blades in the annular diffuser passage and increasing static pressure of the process fluid therein; and an outlet for receiving process fluid discharged from the annular diffuser passage; introducing a first process fluid, comprising natural gas, having a molecular weight (MW) of 12-20 into the inlet passage of the compressor at an inlet pressure (P₁); driving the centrifugal impeller at a rotational speed (N), so that the trailing edges of the respective impeller blades achieve a rotational velocity (U₂) of greater than or equal to 1400 feet/second, imparting kinetic energy into the first process fluid; receiving the first process fluid discharged by the trailing edges of centrifugal impeller in the annular diffuser passage, converting the kinetic energy imparted in the first process fluid by centrifugal impeller into a pressure increase; and discharging the first process fluid from the annular diffuser passage at a discharge pressure (P₂) greater than the inlet pressure (P₁) thereof, such that a pressure ratio (r) of the discharge pressure divided by the inlet pressure is greater than or equal to 2.5:1.
 11. The method of claim 10, further comprising combining two of the single-stage, centrifugal compressors in series, with the outlet of the first stage coupled to the inlet of the second stage, the stages in combination receiving the first process fluid in the inlet of the first stage thereof, and discharging the first process fluid out of the outlet of the second stage at a throughput pressure ratio (r) of greater than or equal to 5:1.
 12. The method of claim 11, comprising configuring and assembling a multi-stage compressor, having 3 or more of the single-stage, centrifugal compressors in series, the stages in combination configured to impart sequentially within the assembled multi-stage compressor a throughput pressure ratio (r) of greater than or equal to 10:1 on the first process fluid.
 13. A method for compressing process fluids, comprising: providing a single-stage, centrifugal compressor with a compressor casing having therein: an inlet for receipt of a process fluid; a single, unshrouded, rotatable, centrifugal impeller defining a plurality of impeller blades for imparting kinetic energy into the process fluid, each of the respective impeller blades having a leading edge for receiving process fluid from the inlet and a trailing edge for discharging process fluid therefrom; a diffuser, defining an annular diffuser passage, for receiving the process fluid discharged from the respective trailing edges of the impeller blades in the annular diffuser passage and increasing static pressure of the process fluid therein; and an outlet for receiving process fluid discharged from the annular diffuser passage; introducing a first process fluid having a molecular weight (MW) into the inlet passage of the compressor at an inlet pressure (P₁); driving the centrifugal impeller at a rotational speed (N), so that the trailing edges of the respective impeller blades achieve a rotational velocity (U₂) of greater than or equal to 1400 feet/second, imparting kinetic energy into the first process fluid; receiving the first process fluid discharged by the trailing edges of centrifugal impeller in the annular diffuser passage, converting the kinetic energy imparted in the first process fluid by centrifugal impeller into a pressure increase; and discharging the first process fluid from the annular diffuser passage at a discharge pressure (P₂) greater than the inlet pressure (P₁) thereof, such that a pressure ratio (r) of the discharge pressure divided by the inlet pressure is: at least 5:1, where the first process fluid has a molecular weight of 24-27.99; or at least 4:1 where the first process fluid has a molecular weight of 20-24; or at least 3:1 where the first process fluid has a molecular weight of 16-20; or at least 2.5:1 where the first process fluid has a molecular weight of 10-16; or at least 2:1 where the first process fluid has a molecular weight less than
 10. 14. The method of claim 13, further comprising combining two of the single-stage, centrifugal compressors in series, with the outlet of the first stage coupled to the inlet of the second stage, the stages in combination receiving the first process fluid in the inlet of the first stage thereof, and discharging the first process fluid out of the outlet of the second stage at a throughput pressure ratio (r) of: at least 10:1, where the first process fluid has a molecular weight of 24-27.99; or at least 8:1 where the first process fluid has a molecular weight of 20-24; or at least 6:1 where the first process fluid has a molecular weight of 16-20; or at least 5:1 where the first process fluid has a molecular weight of 10-16; or at least 4:1 where the first process fluid has a molecular weight less than
 10. 15. The method of claim 14, comprising configuring and assembling a multi-stage compressor, having 3 or more of the single-stage, centrifugal compressors in series, the stages in combination configured to impart sequentially within the assembled multi-stage compressor a throughput pressure ratio (r) of greater than or equal to 10:1 on the selected process fluid having a molecular weight of 2.0-27.99, the selected process fluid comprising natural gas. 